Cranktrain Design Optimized for Reducing Weight and Friction
Supported by Coupled CAE Tools
ABSTRACT
The OEMs are currently being faced with a variety of contradictions in terms of market demands and governmental
regulations, which force them to develop concepts that assure high fuel economy, low exhaust emissions and high
specific power. Because mechanical losses can account for up to 10% of the fuel energy used, the key to achieving these
specific customer/government demands is to enhance the mechanical performance of the engine by decreasing friction
and through the development of lightweight engine parts. Reaching these goals requires that each of these components
be checked carefully to ensure that the designs are fully optimized. Numerous studies have shown that further
optimization of the cranktrain parts is still possible, especially for the crankshaft. Advanced calculation approaches are the
only means to completely exploit the potential development of the crankshaft and to ensure the component layout is within
design limits. The purpose of this study is to illustrate a calculation procedure that highlights the borderline design of
crankshafts under the consideration of the optimization objectives. The results show that even a production crankshaft
contains enough durability potential to optimize its friction and weight, while resulting in minimal or no changes to the
cranktrain and engine block.
INTRODUCTION
Rising fuel prices, competition and legal demands within the last decade has turned the focus of increasing expectations
of engines and as a result increased concentration on optimizing the crankshaft for mass and friction reduction. The
crankshaft is the primary rotating part within an engine. The critical nature of the crankshaft requires R&D engineers to
carefully consider their calculations and optimization plans because this component has to demonstrate durability
throughout the entire engine life cycle. Crankshaft durability creates a variety of development issues, which include mass
reduction, friction reduction, noise reduction and stress reduction according to the defined requirements for the crankshaft
under development. Considering that many of these issues contradict on another, the goal of the optimization has to be
predefined.
The sources of engine mechanical losses are the moving contact surfaces of the components, such as the piston group,
connecting rod, crankshaft and valvetrain. Reducing these losses can only occur through modifications to these
components in order to minimize the arising friction. The frictional losses of the crankshaft are primarily caused by
tribological interactions in the main bearing and pin bearing regions. Figure 1 shows that the crankshaft contributions to
friction mean effective pressure are approximately 20% to 40% (including the crankshaft pin bearing), depending on the
engine speed and oil temperature. The friction amplitude of the components depends on the dimensions of the running
surfaces. Consequently, the greater the contact area is the higher the friction mean effective pressure. Therefore,
decreasing the rotating contact area of the crankshaft effectively reduces the losses. The contact areas can be reduced
with smaller bearing diameters and/or shorter bearing lengths. Exponential benefits to reducing the area are received by
decreasing the diameter, while decreasing the length only provides linear benefits.
Figure 1: Contribution of the Components in the Engine Friction Mean Effective Pressure [1]
(Crankshaft Pin Bearing Friction is Illustrated under the Piston Group/Con-Rod)
Designing lightweight engine components results in another important method of improving fuel economy and engine
dynamic behavior. Additional advantages can be achieved by reducing the mass and inertia moment of the moving
engine parts, which results in decreasing the power requirement for the acceleration phase of the moving components. A
low moment of inertia of the crankshafts has a significant impact on the dynamic response of engines to speed demand.
1
cylinder block
21%
other parts
42%
crankshaft
11%
cylinder head
11%
oil pan
2%
camshaft
2%
Flywheel 5%
piston conrod group
4%
Figure 2: Contribution of the Components in Engine Bulk Weight
The crankshaft’s contribution to the total weight of the engine is shown in Figure 2. Numerous studies have shown that
crankshafts possess a tremendous potential for mass optimization. In 2006, the crankshaft manufacturer Intermet Corp.
showed that a 50% weight reduction can be accomplished with crankshafts [2]. A variety of methods exist for reducing
crankshaft weight, which includes optimizing the web design, smaller bearing diameter and hollow crankshaft designs.
Recently, research and innovation in cast iron materials point to such examples as the use of austenitic ductile iron and
the application of hollow crankshafts. The use of hollow designs for mass reduction leads to a decrease in stiffness, which
requires an intensive investigation into the durability of the crankshaft.
The purpose of this particular study is to illustrate the optimization potential of a cast iron crankshaft for an inline 4cylinder engine.
3D DURABILITY ANALYSIS OF CRANKSHAFTS
A vital requirement in order to optimize a system is being aware of the potential which is available for the strived trade off.
This means the durability calculation of the crankshaft plays an essential role for the optimization, since the modified
crankshaft acc. to weight and friction targets have to endure as long as the crankshaft, without modifications. In order to
have a reliable conclusion on the safety of the crankshaft design, durability analysis of the crankshaft has to be applied
under real engine operation cycles and with actual material behaviors.
simulation pre processing
data flow
1 FEA mesh generation
FEA mesh of flexible component
2 FE Software modal analysis
component mode shapes
3 ADAMS/Engine powered by FEV
3D full dynamic simulation
post processing
scale factors
4 FE Post processor
3D stress data in time domain
5 commercial Fatigue Software
safety factors
FEA software
MBS software
output
Figure 3: Coupled Dynamic Simulation Process for Crankshaft Durability Assessment [3]
2
The highly profound level of such simulations, combined with the contradicting goals to reach higher loads and reduce
engine mass or friction without any endurance problems, make multi-purpose simulation tools indispensable for these
durability calculations. In recent years the combination of multi-body dynamics and structural dynamics using a modal
condensed FEA structure in MBS simulations has become a standard procedure. The coupling of fatigue analysis to this
process chain leads to a better overview and an efficient methodology.
The process for simulation of fatigue strength of the crankshaft can be subdivided into the following steps as illustrated in
Figure 3:
1. Pre processing: Eigen mode FEA of crankshaft assembly and setup of cranktrain multi-body system
2. Simulation: Calculation with the fully coupled dynamic MBS model
3. Post processing: Evaluation of dynamic cranktrain behavior and durability and/or lifetime prediction
PRE PROCESSING - The model is set up in two parallel steps. On one hand, the crankshaft is discretized in finite
elements in order to create a flexible structure for modal analysis. On the other hand, an MBS model of the cranktrain is
generated in FEV Virtual Engine for multi-body calculations. Once both processes are accomplished, the modal
condensed crankshaft is coupled with the prepared multi-body model. In this way a hybrid model is created.
SIMULATION – The advanced crankshaft simulation approach of FEV Virtual Engine is based on modal condensation,
which combines natural modes with modes driven by specific nodal degrees of freedom. So the primary result of the MBS
Analysis is modal scale factors (one per mode and time step). A transient rpm sweep is simulated in order to record the
resonance points in the speed range. Then a steady-state analysis is applied for the specified engine speeds in order to
calculate the modal scale factors. This is utilized for post processing to determine the safety factors.
POST PROCESSING - The core of the fatigue strength analysis is the generation of the 3D safety factor distribution on
the FEA mesh of the crankshaft. In this way the fatigue durability of the component can be revealed. To determine the
safety factors in the first step, scale factors are multiplied with the mode shapes and modal stresses and are superposed
for each time step. The result is the time history of structural deformation and stress. So with a resolution of one degree of
crank angle, there are 720 deformation and stress matrices of the crankshaft assembly for each simulated speed point.
From the 720 stress matrices, distributions of mean stress and stress amplitudes can be calculated by means of proven
procedures, as e. g. the critical cutting plane procedure. These distributions are then used to determine the distribution of
safety factors, using locally variable fatigue strength graphs [4].
CRANKSHAFT OPTIMIZATION
A gasoline four-cylinder engine with a cast iron crankshaft was selected to investigate the applicability of the weight and
friction optimization. The selected crankshaft is from a naturally aspirated gasoline engine, which lies slightly under the
average line in concern of both friction and mass characteristics.
Figure 4: FE Model of the Selected Cranktrain
In order to optimize the crankshaft, the potentials of the current design have to be revealed. The method illustrated in
Figure 3 is utilized to calculate the fatigue behavior of the crankshaft. The crankshaft is manually meshed in order to have
high mesh intensity in the fillet radii to obtain proper results and a coarse mesh in the webs, counterweights and flywheel
to save elements in these zones. Crankshaft pin and journal fillets are meshed with at least four equidistant layers
through the depth in the radius. Oil bore outlets are also discretized fine in order to model the real contour. The final
crankshaft finite element model has approximately 1.8Mio. nodes with second order linear tetrahedron elements, as
shown in Figure 4.
3
Figure 5: Coupled MBS Model of the Cranktrain
After-modal analysis acc. to Craig/Bampton, the finite element crankshaft model (flexible model) is coupled with a multibody system model, Figure 5. This kind of model can be named as E-MBS (Elastic Multi-Body System), since a
conventional MBS model does not contain any deformable body. The dynamic loading situation is simulated with FEV
Virtual Engine in order to determine the scale factors for each mode throughout the speed range from 1000 rpm to 7000
rpm with a 250 rpm step size. Since the load history is defined, a fatigue analysis is applied for each speed step. After the
superposition of the 3D stress data in time domain, a safety factor assessment is carried out with the commercial fatigue
software FEMFAT.
Minimum safety factors for the base crankshaft version were depicted throughout the engine speed range. Based on the
calculations, it can be predicted that the crankshaft is durable throughout the entire engine life within the speed range
1000 rpm to 7000 rpm under full load. If the progression of the safety factor curve is closely analyzed, a clear distance
between the safety limit and calculated safety factors is observed up to 6750 rpm. Therefore, investigations to reach to an
optimized crankshaft will be conducted relative to the base design.
Once the potential of the crankshaft is evident, a strategy can be selected to exploit the room which is available for
optimization. There are different ways available to utilize the existing potential:
•
•
•
•
Stress reduction
Noise reduction
Mass reduction
Friction reduction
If the 3D durability analysis shows that the crankshaft is under the safety limit, there is a clear need for stress reduction on
the crankshaft. This can be achieved either by local design changes, which alters the force flow in the crankshaft so that
the critical areas are relieved or by global design changes which increase the stiffness of the whole crankshaft.
Another case is that there is a need for stress reduction, even though the crankshaft exhibits durability throughout the
entire engine lifetime based on 3D analysis. This would be the case if the engine has to be certified according to some
classifications defined by associations or insurance companies e.g. CIMAC, the International Council on Combustion
engines, or DNV, Det Norske Veritas etc., which is mostly the case in marine applications. Since such layout approaches
are more conservative than the illustrated methodology above, a stress reduction may be desirable.
Improvement on the crankshaft acoustic behavior is not necessarily a contradicting phenomenon with the durability of the
crankshaft. Due to high acoustical interactions between the separate engine components, simple judgments are not
generally accepted. But stiffness increase contributes mainly to noise reduction of the crankshaft. Therefore, acoustic
optimization correlates mainly with the crankshaft durability.
Utilization of the potential for weight optimization is the most common strategy. In most of the cases the reduction of the
mass brings along the durability problems due to the loss of static stiffness of the crankshaft. There are two ways to cope
with this problem. The first option is the optimization of the crankshaft design, which leads to the same static stiffness
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despite a mass decrease. The second way is to calculate the resonance regions of the crankshaft and remove material
accordingly to reach a convenient natural frequency, which would mean a higher dynamic stiffness of the crankshaft.
Reduction of the mechanical losses is another important aspect which could be selected as a strategy to make use of the
potential. In order to reduce the friction, the rotating contact surfaces should be as minimal as possible, which means
smaller bearing diameters for the crankshaft, since the bearing width should be long enough to build a load carrying oil
film. This fact contradicts with the durability because the stiffness is highly dependent on the overlapping of the
crankshaft. Since the mass decrease due to diameter reduction is marginal, an increase of the dynamical stiffness cannot
be realized in this case. Therefore, a clear trade off has to be met in case of the reduction of the crankshaft bearing
diameters. In this paper two scenarios are investigated in order to exploit the room for optimization, based on the
durability behavior data gathered.
The first scenario uses the potentials to reduce the friction generated by the crankshaft. The most effective way to reduce
crankshaft friction is to decrease the bearing diameters. In order to comprehend the individual influences, reduction of pin
bearing and journal bearing diameter are applied separately as shown in Figure 6. A unit dimension reduction of 1 mm is
selected so that the effect is easily quantifiable. This is equal to approximately a 2% decrease in pin and journal bearing
diameters. There is no counterweight modification undertaken, since the change of mass in pin bearings is marginal.
After the safety analysis of the variants “reduced pin diameter” and “reduced journal diameter”, the safety factors of these
variants will be compared with the base design. If the variants do not fulfill the durability requirements, it will be concluded
that the crankshaft cannot be optimized in terms of friction. But if the crankshaft is still durable after the dimension
changes, the benefit of the variants can be quantified in terms of reduction in Friction Mean Effective Pressure (FMEP). In
this way, the trade off between the durability and friction can be revealed for an existing crankshaft of an inline 4 engine.
Figure 6: Crankshaft with Reduced Bearing Diameters
The second scenario uses the potentials to reduce the mass of the crankshaft. In order to reduce the weight of the
crankshaft, a strategy is selected which should not cause any change in the dimensions of the cranktrain. In this way the
change in mass can be compared properly, without taking other cranktrain components into consideration. This goal can
be reached either by changing the web design or by applying a hollow design of the crankshaft bearings. The latter
solution is selected for the investigations. Similar to the procedure in the first scenario, the hollow design of the pin and
journal bearings are investigated separately. There are many different examples and patents in literature how a hollow
crankshaft may be designed. For the investigations, a very conventional hollow design is selected and realized and
depicted in Figure 7. Hollow rooms of approximately 50% of the pin and journal diameters are created for the respective
variants.
After the safety analysis of the variants “hollow pin design”, “hollow pin design with modified counterweights” and “hollow
journal design”, the fatigue behaviors will be compared with the base design. An identical decision method will be followed
as in the first scenario to interpret the benefit of the variants. In this case, the mass decrease will be quantified and
correlated against the base design considering the changes in safety factors. The mean and maximum bearing forces will
also be compared with other variants.
Figure 7: Hollow design of the crankshaft
5
The reduction of crank pin mass enables a further reduction in counterweight mass, since the balancing ratio is to be kept
constant. The “m*r” value of each crank of the base design is taken as reference value for the mass balancing. The
counterweights of the variant “hollow pin design with modified counterweights” are modified to reach the same mass
balancing ratio as the base design.
FRICTION REDUCED CRANKSHAFT
The engine’s primary source of friction is the piston/cylinder system followed by the crankshaft, as illustrated in Figure 1.
Crankshaft friction originates mainly from the pin and journal bearings and slightly from the rotary shaft seals. Therefore,
reducing the friction in crankshaft is only possible with systematic modifications on the pin and journal contact surfaces.
An alternative way to reduce the friction in crankshaft bearings is changing the bearing concept by using roller bearings
instead of plain bearings. By substituting plain bearings with roller bearings, a high friction potential of crankshafts up to
55% has been proven by test bench measurements applied by FEV [5]. In addition, the oil volume flow can be reduced e
using roller bearings, a fact which also contributes to friction reduction. According to the solution developed by FEV,
which has been implemented in a demonstration vehicle, the replacement of the mounting of the crankshaft and
connecting rods to roller bearings, together with the adjusted oil pump at a part load operating point of 2000rpm at 90°C,
results in a reduction of engine friction loss of approximately 25% [5]. Nevertheless, there are some drawbacks that may
limit the implementation of such a solution. The NVH behavior of the crankshaft will be highly influenced of this
modification, which makes a detailed investigation essential for each separate case.
If the aim is to reduce the friction losses with marginal changes on crankshaft, the most efficient way is to reduce the
crankshaft bearing diameters in order to minimize the rotating contact surfaces. For the investigations a strategy to
decrease the pin and journal bearing diameters was selected.
During the investigations, two separate variants were applied in order to reveal the individual effects of pin bearing and
journal bearing diameter reduction. The resulting safety factors of both variations were compared with the base variant in
order to monitor the change in fatigue durability for the selected engine speeds.
Safety Factor [-]
Basis
Reduced Pin ø
Reduced Journal ø
Safety Limit
1000
2000
3000
4000
5000
6000
7000
Engine Speed [1/min]
Figure 8: Comparison of the Minimum Safety Factors
The comparison of the minimum safety factors of the variants with the base design is depicted in figure 8. It has to be
noted that the position of the minimum safety factors changes throughout the speed range. However, the progression of
the curves shows that there is a minimal change in natural frequency of the variants. This proves that the reduction in
stiffness is in the same range as the mass inertia decrease.
Figure 9: The Position of the Critical Crankshaft Sections
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Reduction of the journal diameter does not lead to a significant change in safety factors until 6000 rpm. Nevertheless,
beginning from this speed a steep reduction is observed, leading to a durability problem at 7000 rpm. On the other hand,
reduction of pin bearing diameter exhibits lower safety factors up to 6000 rpm, hence leading to no fatigue problems.
However, in the high speed range the durability behavior shows nearly no difference compared with the base version. In
order to understand this interesting fact, a closer observation of the individual positions, depicted in Figure 9, is
necessary.
The minimum safety factor is not observed in the same position throughout the entire speed range, since the most critical
point is changing, depending on the eigen (natural) modes of the crankshaft. Alteration of the compression and tensile
stresses play a big role on the shifting of the most critical position. It was noted that up to 6000 rpm, pin bearing fillets
exhibit the minimum safety factors; however, above this speed the most critical position shifts to the 4th main bearing. This
proves why the variant with reduced journal diameter experiences a sharp decrease in high speeds.
When the safety factor trends of 2nd and 4th pin fillets are observed, the variant with reduced pin diameter has lower safety
over the entire speed range. This fact is reasonable, since decrease of the pin diameter leads to a higher local stiffness
reduction at the pin bearings. The safety factor progression of the 4th main bearing experiences a rapid decrease
beginning from 5000 rpm and the variant with journal diameter reduction exhibits lower safety factors than the variant with
pin bearing reduction. The journal bearing variant is even under the safety limit at 7000 rpm. These two facts show that it
is not possible to generalize whether pin bearing or journal bearing diameter reduction leads to fatigue problems. Since
the most critical deformation mode is changing, a prediction is only possible if the base variant is simulated in detail
analysis. If the location of the most critical point is evident, the effect of diameter change can be predicted effectively.
Once the effects of the variations on fatigue life are evident, the benefit of the modifications can be investigated. In this
case the profits in friction behavior have to be searched. An appropriate way to compare the friction behavior is to
examine the change in mean friction torque resulting from crankshaft main bearings. A reduction in mean friction torque
would lead to approximately the same percentile decrease in the crankshaft portion of the Friction Mean Effective
Pressure (FMEP). Therefore, a deviation of this value will affect the FMEP amplitude.
Table 1: Change in Friction Torque at Main Bearing in Comparison to Base Variant
Engine Speed
1000
2000
3000
4000
5000
6000
7000
Change in Friction Torque
Red. Ø Pin
Red. Ø Journal
0%
-1%
-1%
-1%
-1%
-2%
-3%
-5%
-5%
-5%
-5%
-5%
-5%
-5%
In table 1, the deviation of mean friction torque in comparison with the base design is depicted. A 2% reduction of journal
bearing diameter leads to an approximate 5% decrease in friction torque throughout the entire engine speed, whereas a
2% reduction in pin bearing diameter shows more speed dependency and a progressive decrease in the friction torque
resulting with a 3% reduction at maximum engine speed. In order to quantify the effect of this reduction in friction, the
change in FMEP should be determined. Considering the fact that crankshaft main bearing contributes approximately 20%
to FMEP acc. to Figure 1, a decrease of above 1% in FMEP can be expected, due to a reduction of approximately 2% in
the main bearing diameter, which is in this case equal to 1 mm main bearing diameter reduction.
LIGHTWEIGHT CRANKSHAFT
The crankshaft is one of the bulkiest components in an engine, following the engine block and head. As illustrated in
Figure 2, the crankshaft has approximately a 10% contribution to entire engine weight. Therefore, the mass reduction of
crankshafts plays an important role for the light weight of the whole engine. Since the crankshaft is not a stationary
component, the decision where to remove the material has to be determined carefully. The same amount of material
removal from different regions of the crankshaft results in outstanding changes in deviation moment, mass balancing ratio
and natural frequency of the cranktrain. Hence, different solutions have to be reviewed to select the mass reduction
methodology.
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One way to reduce the crankshaft mass is to decrease the balancing ratio in order to reduce the mass of the
counterweights. Although this is an effective solution, it affects crankshaft dynamics negatively and leads to an increase in
bearing forces. Therefore, in this study no single modification in counterweights will be undertaken.
Another way to reduce the mass is modification of crank web shape, so that same mass of inertia can be obtained with
less mass, resulting with unchanged balancing ratio. A structural optimization approach with the aid of finite element
method would be helpful to reach an optimal crank web design. Also the manufacturing process of the crankshaft has to
be taken into consideration, so that the optimized crank web shape is realizable. For elevated speeds the durability of
such a crank web can be critical; therefore, for high speed concepts a fatigue analysis would be necessary.
For the following study, a hollow design concept is selected for the variations. An outstanding reason is that no
modification should be applied to the other parts of the cranktrain. Another reason is that the crankshaft is to be produced
with a casting process; so the hollow design is realizable with conventional production procedures.
In the frame of the investigations, three separate variants are applied in order to reveal the individual effects of hollow
design of crankshaft pin bearing and journal bearing. For the hollow pin design, two variations are applied. For both of the
variations, 50% of the pin bearing inner diameter is removed. In the variant “hollow pin with modified counterweights”, a
further modification of the counterweights is applied to obtain the same balancing ratio as before. In this case the
necessary amount of mass from counterweights is removed in order to reach same “m*r” value for each individual crank.
For the hollow journal design, approximately 50% of the main bearing inner diameter is removed as illustrated in Figure 7.
No further changes are applied for this variant.
The comparison of the minimum safety factors of the variants with the base design is depicted in Figure 8. The hollow
journal design exhibits even higher safety than a base variant up to 5750 rpm. This shows that the dynamic stiffness is
higher up to this speed. After 6000 rpm, the most critical position is shifted to main bearing fillet; therefore, a rapid
reduction in safety factors is observed for hollow journal variant. Since mass reduction in main bearings does not lead to a
significant change in mass inertia moment, high speed ranges are more critical for this design because of the stiffness
reduction resulting from the reduced overlap. Furthermore, mass removal in the 5th bearing leads a reduction of stiffness
in flywheel flange area. Therefore, the last main bearing is objected to increased gyroscopic loads in higher speeds. As in
the previous friction variant of “reduced journal diameter” (see Figure 10), the variant “hollow journal design” encounters
durability problems at 7000 rpm.
It can is observed from the table 2 that the natural frequency, ωE, is reduced by 1% in the case of the hollow journal. It
means stiffness, cT, decrease is higher than moment of inertia, θΤ , decrease acc. to equation 1. Whereas, a higher mass
inertia decrease than stiffness decrease is reached for the variant “Hollow pin with modified counterweights”, leading a
natural frequency increase of 4.4%. No specific change is observable in the natural frequency of the “hollow pin” variant in
comparison with the base variant.
Safety Factor [-]
Basis
Hollow Pin
Hollow Pin w modif. counter weights
Hollow Journal
Safety Limit
1000
2000
3000
4000
5000
6000
7000
Engine Speed [1/min]
Figure 10: Comparison of the Minimum Safety Factors
8
None of the hollow pin variants encounter safety problems within the observed speed range. The main difference
between the hollow pin variants is the deviation in natural frequency. But the differentiation in safety factor progression
cannot be explained only with this single effect on crankshaft dynamics. At 3000 rpm, which can be named as the first
resonance region, there is no significant difference between the hollow pin variants, since the stiffness matrices of these
variants are quite similar. Whereas up to 6500 rpm, a second resonance region begins and a clear difference between the
fatigue behaviors can be observed. Apparently, reduced mass inertia benefits the crankshaft dynamics beyond the
resonance regions for the “hollow pin with modified counterweights” variant.
ωE =
cT
Equation: 1
θT
Table 2: Change in Natural Frequency and Mass in Comparison to Base Variant
Variants
Hollow Pin
Hollow Pin w.
modified c. w.
Hollow Journal
Change in
Natural Frequency
~0%
Mass
-3.2 %
+4.4 %
-6.2 %
-1.0 %
-3.7 %
When the safety factor progressions at the 4th pin bearing fillet on flywheel side are observed and compared with the
minimum safety factors curves in figure 10, it can be seen that the most critical point for hollow pin variants is the 4th pin
bearing right fillet up to 6500 rpm. Whereas, for the hollow journal variant the most critical point is shifting the position
between 1st, 2nd, 3rd pin bearing and 4th, 5th main bearing fillets, as illustrated in figure 11. It shows that the most critical
deformation mode of the crankshaft is changing over the speed range in the case of hollow journal variant.
Safety Factor [-]
Pin1L / Hollow Journal
Pin2R / Hollow Journal
Main4R / Hollow Journal
Pin4R / Hollow Journal
Main5L / Hollow Journal
Safety Limit
1000
2000
3000
4000
5000
6000
7000
Engine Speed [1/min]
Figure 11: Comparison of the Safety Factors of “Hollow Journal” at Different Fillet Positions
The hollow design of the last main bearing illustrated in figure 7 caused a stiffness reduction in this region. Considering
the fact that the flywheel has a significant effect in this area, the 5th main bearing became critical in the high speed region
beginning from 5500 rpm, as shown in figure 11.
The reduction of the mass, due to the variations, is depicted in table 2. For “hollow pin bearing with modified
counterweights”, a reduction of more than 6% is acquired, without having fatigue problems. This is equal to approximately
a 0.6% mass reduction for entire engine. A mixed variant, which includes hollow pin and hollow journals at the same time,
would lead to a mass decrease of approximately 10%, which is a considerable reduction for the whole engine by more
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than 1%. A more aggressive weight saving can be realized, if the crank web designs are modified. A systematical
optimization would lead further mass decrease with minimal influence on crankshaft fatigue behavior.
COMPARISON OF THE FRICTION REDUCED AND LIGHTWEIGHT CRANKSHAFT RESULTS
In the previous sections, the investigated variants were introduced, the changes in durability behavior are illustrated and
the resulting benefits are quantified. In figure 12, the safety factor curves of all variants and the base design are illustrated
in the same diagram. All of the variants can be realized up to 6750 rpm, without fatigue problems. At a maximum speed of
7000 rpm, both journal bearing variants lie slightly under the required safety limit.
Safety Factor [-]
Basis
Reduced Pin ø
Reduced Journal ø
Hollow Pin
Hollow Pin w modif. counter weights
Hollow Journal
Safety Limit
1000
2000
3000
4000
5000
6000
7000
Engine Speed [1/min]
Figure 12 – Comparison of the Minimum Safety Factors of the Friction and Weight Reduction Variants
Since the crankshaft interacts with other components, such as the engine block and crankcase, the influence of the
variants on these parts should be also investigated. In Table 3, the main bearing forces on a selected bearing at 7000
rpm are depicted. Since the change in main bearing forces at low speeds is insignificant, this speed is selected to show
the highest deviations resulting mainly from the difference in the mass matrix.
For friction variants where change in mass is insignificant, a neglectable change in bearing forces is observed. For hollow
pin variants, higher changes are noticed. In the case of the variant “hollow pin with modified counterweights,” where the
balancing ratio is kept constant, maximum force decreases because the total mass is reduced more than 6%. For the
variant “hollow pin,” where the balancing ratio increases since the counterweights are not modified, a significant increase
is observed in maximum bearing force. A reduction of mean force may be expected for the latter variant, due to the
balancing ratio increase, but apparently the worsening dynamical behavior of the crankshaft causes higher bearing loads.
The curve of the variant “hollow pin” in Figure 12 also shows the reduced safety factors of the hollow pin crankshaft
throughout the speed range.
Table 3: Change in Main Bearing Forces in Comparison to Base Variant at 7000 rpm.
Variants
Hollow Pin
Hollow Pin w.
modified c. w.
Hollow Journal
Red. Ø Pin
Red. Ø Journal
Change in Main Bearing
Mean Force
Max. Force
+3.6%
+8.5%
+3.0%
-3.6%
~ 0%
+1.5%
-0.7%
-2.8%
~0%
-1.0%
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A further comparison of minimum oil film thickness at the bearing can be applied to investigate the safety of the bearings.
Another comparison may be to show the difference in speed fluctuation in free end or the deflection angle difference to
investigate the influences on secondary drive dynamics.
SUMMARY AND CONCLUSION
This study focused on the optimization possibilities in the cranktrain, with specific emphasis on the crankshaft. A
crankshaft made from spheroidal graphite cast iron was selected to utilize the form flexibility of the casting process and
the cost benefits of using iron.
First, the safety factors of the crankshaft were investigated to establish its potential for being exploited. The methodology
outlined in Figure 3 was used for calculating the safety factors of the crankshaft. This methodology allows the observation
of the dynamical behavior of the crankshaft in the time and frequency domain, either in stationary or transient cases.
Once the potential is clear, the variations are applied and compared with the base design. The benefits gained with the
various modifications are then document in tables.
The results illustrate that even a production crankshaft possesses enough potential for further optimization in friction and
weight reduction. According to the outputs of the simulations, the goals defined by the optimization are realizable, without
encountering fatigue problems. Additional reductions in bearing diameters or larger inner diameters would lead to a loss
in durability, which can cause the crankshaft to operate under the safety limit, even at lower speeds.
Since the gains are not marginal and the need for modification of other cranktrain components is minor, any of the
variations investigated in this study can be applied, even in a late phase of a development cycle. However, the purpose of
this development should be clear, so that the method of exploiting the component’s potential can be selected.
The variations applied in this study are unit changes, such as a 2% diameter reduction or a removal of 50% of the inner
diameter. This is selected to comprehend the influences of unit changes on crankshaft dynamics in detail. Whereas, an
optimization algorithm that searches for the required diameter reduction or inner diameter, which has a minimum safety
factor within the safety limit, can fully exploit the potentials that a crankshaft possesses. However, due to time and
hardware constraints this optimization is only realizable for simple crankshaft models. Since simple crankshaft models are
not reliable for borderline component designs, an approach of this nature cannot guarantee the utilization of the potentials
of a crankshaft. Future developments to reach a weight and friction optimized crankshaft can be summarized as solving
this paradox.
REFERENCES
[1]
[2]
[3]
Pischinger, S., “Internal Combustion Engines”, 2008.
Druschitz, A. P., Fitzgerald, D. C., Hoegfeldt, I., “Lightweight Crankshafts”, SAE-Paper, 2006-01-00162006, 2006.
Ortjohann, T., Rebbert, M., Maassen, F., Robers, M., “3D-Durability Analysis of Crankshafts via Coupled Dynamic
Simulation including Modal Reduction”, SAE, 2006 06B-363, 2006.
[4] Ortjohann, T., Rebbert, M., Maassen, F., Robers, M., “Kurbelwellensimulation”, MTZ, 05-2006, 2006.
[5] Schwaderlapp, M., Maassen, F., Lang, O., Koerfer, T, Hermsen, F.-G., “Friction Reduction – The Mechanical
Contribution to Fuel Economy”, ATA Conference, 2006.
CONTACT
FEV Motorentechnik GmbH
Neuenhofstr. 181
52078 Aachen
[email protected]
DEFINITIONS, ACRONYMS, ABBREVIATIONS
ωE = Natural frequency
cT = Stiffness
θΤ = Moment of inertia
m = Crank mass
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r = Center of gravity position of crank
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